Compressor

ABSTRACT

A turbocharger unit for an internal combustion engine having at least one exhaust line for the evacuation of exhaust gases from the combustion chamber of the engine and at least one inlet line for the supply of air to the combustion chamber is provided. The turbocharger unit includes a turbine which interacts with a compressor in order to extract energy from the exhaust-gas flow of the engine and pressurize the inlet air of the engine. The compressor is of the radial type and is provided with an impeller having backswept blades in which the blade angle between an imaginary extension of the center line of the blade between root section and tip section in the direction of the outlet tangent and a line connecting the center axis of the impeller to the outer tip of the blade is at least about 40 degrees. The ratio between the inlet diameter of the impeller and its outlet diameter lies within the range 0.50-0.62. The compressor diffuser is provided with blades, the length of which has a ratio to the distance between them, along the periphery in the blade inlet, within the range 0.7-1.5.

The present invention relates to a compressor in a turbocharger unit foran internal combustion engine.

The state of the art relating to turbocharger systems for thesupercharging of diesel-type internal combustion engines, preferably forheavy-duty vehicles, usually comprises a single-stage compressor whichis driven by a single-stage turbine, both of the radial type.

Superchargers suitable for a diesel motor of 6 to 20 liter cylindercapacity normally have an efficiency, under stationary conditions, ofbetween 50% and 60% (ηcompressor*imechanics*ηturbine). In present-daydiesel engines, the benefit from good efficiency is lower than forfuture engines, which will require higher boost pressure. Examples ofsystems which raise the supercharging requirement arc exhaust-gasrecirculation for lower nitrogen oxide emissions, or systems involvingvariable control of inlet valves.

Higher-efficiency turbocharger systems offer an increased chance ofmeeting future requirements for environmentally friendly and fuel-leanengines. Hitherto, environmental requirements for diesel engines haveusually resulted in impaired engine efficiency, which has thus meantthat poorer use is made of the energy content of the fuel.

The compressor in a turbocharger unit consists of or comprises a bladeddisc, an impeller, mounted on a shaft for rotation in a compressorhousing. The compressor housing is constituted by a stationary shroud atshort distance from the blades, typically less than 0.5 mm, a radiallydiffusing duct section, a so-called diffuser, and a volute outlet. Uponrotation of the shaft, gas, for example air, is sucked into theimpeller, whereupon the static pressure level and flow velocityincrease. During passage through the diffuser, some of the velocityenergy of the gas is converted into a further static pressure increase,whereafter the gas leaves the compressor via the outlet. The voluteshape of the outlet, as a result of its divergent duct shape, alsocontributes to a certain increase in the static pressure level.

The efficiency (i.e. isentropic efficiency), pressure build-up andstable working range (flow interval for a given rotation speed) of acompressor are substantially determined by its dimensions and bladeconfiguration. For present-day compressors, the ratio between the inletdiameter of the impeller and its outlet diameter lies usually within therange 0.63-0.70. Present-day impellers are usually provided withbackswept blades, in which the blade angle βb2, between an imaginaryextension of the center line of the blade between root section and tipsection in the direction of the outlet tangent and a line connecting thecenter axis of the impeller to the outer tip of the blade, lies withinthe range 25-40°.

A drawback with increasing the blade angle βb2 of the compressor is thatthe peripheral velocity and thus the tensions in the impeller, above allin the hub of the disc and the outlet of the blade, increase for thesame pressure ratio. One method of reducing the tension in the outlet ofthe blade is to incline the blade forward in the direction of rotation,so that the tip of the blade lies in front of its root, i.e. where theblade meets the disc, in the outlet. This angle of inclination ζb2 liesfor present-day compressors typically within the range 0-30°. Forcertain demanding applications, an increase in the blade angle βb2 ofthe compressor may also mean that the impeller must consist of orcomprise a material with higher strength properties. For example, it ispossible to pass from the present-day cast aluminum wheels tosignificantly more expensive forged, worked aluminum wheels or titaniumwheels.

The diffuser in present-day compressors consists of or comprises twoparallel or almost parallel duct walls, in which one duct wallconstitutes a part of the impeller, whilst the other duct wallconstitutes a part of the bearing housing which extends out to thevolute outlet. With a view to improving the efficiency, above all athigh pressure ratios, the diffuser can be provided with aerodynamicallyconfigured blades, so-called guide vanes, which is normally the case forradial compressors in gas turbine applications. The ratio between thelength of the compressor blades (in the direction of flow) and thepitch, i.e. mutual spacing along the periphery, lies within the gasturbine industry typically within the range 3 to 5. This type ofdiffuser with sparsely placed and/or short diffuser blades is sometimesreferred to as a LSA (Low Solidity Airfoil) diffuser.

It is desirable to provide a compressor with improved efficiency,especially at high pressure ratios, i.e. above about 3:1, whilst at thesame time the scope of the working range is not impaired.

In a purpose-built compressor according to an aspect of the invention ina turbocharger unit for an internal combustion engine having at leastone exhaust line for the evacuation of exhaust gases from the combustionchamber of the engine and at least one inlet line for the supply of airto said combustion chamber comprising a turbine which interacts with thecompressor in order to extract energy from the exhaust-gas flow of theengine and pressurize the inlet air of the engine, the compressor is ofthe radial type, having an impeller with diameter ratio Di/Du betweeninlet and outlet within the range 0.50-0.62 and provided with backsweptblades in which the blade angle βb2, between an imaginary extension ofthe center line of the blade between root section and tip section in thedirection of the outlet tangent and a line connecting the center axis ofthe impeller to the outer tip of the blade, is at least about 40°, andin which the diffuser is provided with blades.which have a ratio betweenlength (in the direction of flow) and pitch, i.e. mutual spacing alongthe periphery in the blade inlet, within the range 0.7-1.5.

One advantage of an aspect of the present invention is that dieselengines which demand high pressure ratios, i.e. above 3:1, can bedesigned to attain a higher isentropic efficiency without furtherrestrictions as regards a stable working range. This in turn leads todiminished power dependency and thus reduced specific fuel consumptionfor the engine.

The described compressor can also be used in a two-stage turbo system,which has the advantage that each turbocharger operates with lesspressure increase and thus lower peripheral velocity, so that the use ofmodern materials is facilitated. The higher density of the compressedgas then gives rise to a reduction in the inlet dimension of the secondstage relative to the first stage, so that this compressor, givenoptimal configuration, acquires a diameter ratio approaching the lowerlimit within the range Di/Du=0.50-0.62.

Advantageous illustrative embodiments of the invention emerge from thefollowing description.

BRIEF DESCRIPTION OF THE FIGURES

The invention will be described in greater detail below, with referenceto illustrative embodiments shown in the appended figures, wherein

FIG. 1 shows schematically an internal combustion engine having atwo-stage turbocharger system,

FIG. 2 is a longitudinal section through the two turbocharger stagesincorporated in the turbocharger system,

FIG. 3 shows in partially broken plan view toward the inlet an impellerused in the turbocharger unit according to an aspect of the invention,

FIG. 4 shows a plan view from the side an impeller used in theturbocharger unit,

FIG. 5 shows isentropic efficiency as a function of the mass flow/(massflow maximum) for a compressor according to an aspect of the inventionand the state of the art,

FIG. 6 shows pressure flow as a function of mass flow/ (mass flowmaximum) for a compressor according to an aspect of the invention andthe state of the art,

FIG. 7 shows a work factor for a typical compressor as a function ofblade outlet angle,

FIG. 8 shows a typical compressor efficiency as a function of a specificrotation speed.

DETAILED DESCRIPTION

An aspect of the invention is described used in a two-stagesupercharging system for, in the first place, diesel engines having acylinder capacity between about 6 and about 20 liters, for usepreferably on heavy-duty vehicles such as lorries, buses andconstruction machinery. The supercharging system has the characteristicthat it gives a considerably more effective supercharge than currentsystems. The supercharge is realized in two stages with twoseries-connected compressors of the radial type, with intermediatecooling. The first compressor stage, termed the low-pressure compressor,is driven by a low-pressure turbine of the axial type. The secondcompressor stage, the high-pressure compressor, is driven by ahigh-pressure turbine of the radial type.

FIG. 1 shows an engine block 10 having six engine cylinders 11, whichcommunicate in a conventional manner with an induction manifold 12 andtwo separate exhaust manifolds 13, 14. Each of these two exhaustmanifolds receives exhaust gases from three of the engine cylinders.

The exhaust gases are conducted via separate pipes 15, 16 up to aturbine 17 in a high-pressure turbo unit 18, which comprises acompressor 19 mounted on a common shaft with the turbine 17.

The exhaust gases are conducted onward via a pipe 20 to a turbine 21 ina low-pressure turbo unit 22, which comprises a compressor 23 mounted ona common shaft with the turbine 21. The exhaust gases are finallyconducted onward via a pipe 24 to the exhaust system of the engine,which can comprise units for the after-treatment of exhaust gases.

Filtered inlet air can be taken into the engine via the pipe 25 andconducted to the compressor 23 of the low-pressure turbo unit 22. A pipe26 conducts the inlet air onward via a first charge-air cooler 27 to thecompressor 19 of the high-pressure turbo unit 18. After this two-stageboost with intermediate cooling, the inlet air is conducted onward viathe pipe 28 to a second charge-air cooler 29, whereafter the inlet airreaches the induction manifold 12 via the pipe 30.

The turbocharger system is shown in greater detail in FIG. 2, whichillustrates the double, volute inlets 15, 16 to the high-pressureturbine 17, which each provide half the turbine with gas flow via inletguide vanes 17 a. The high-pressure turbine 17 is of the radial type andis connected to the low-pressure turbine 21 via the intermediate duct20.

The high-pressure turbine 17 is mounted together with the high-pressurecompressor 19 on the shaft 31. The low-pressure turbine 21 iscorrespondingly mounted together with the low-pressure compressor 23 onthe shaft 32.

By combining an impeller with unusually low diameter ratio DIZDn,provided with backswept blades with unusually large blade angle βb2,with a diffuser provided with blades which are unusually short (in thedirection of flow) and/or sparsely placed, a compressor with highefficiency, especially at high pressure ratios, i.e. above 3:1, can beproduced, whilst at the same time the scope of the stable working rangeis not impaired.

Both the low-pressure turbo and the high-pressure turbo have compressorswhich are configured according to the described invention, i.e. impellerwith low diameter ratio between inlet and outlet in combination withhackswept blades with large angle, which will be described below withreference to FIGS. 3 and 4.

From FIG. 3 it can be seen that a blade angle βb2, between an imaginaryextension of the blade 35 along the center line between root section andtip section in the direction of the outlet tangent and a line 36 (indash-dot representation) connecting the center axis of the impeller tothe outer tip of the blade, is at least about 40°. Turbocompressorsavailable on the market have blade angles βb2 between about 25° andabout 40°. The effect of this increase in blade angle lies primarily inthe fact that the impeller with associated turbine rotates at a higherrotation speed for a given pressure ratio. The increase in rotationspeed means that the diameter of the turbine wheel and thus also itspolar mass inertia can be reduced. As a side effect of this, thetransient response of the engine is also improved, since the reducedmass inertia means that the turbine wheel can more easily accelerate toits effective rotation speed range. Moreover, the compressor efficiencyincreases, inter alia by virtue of a diminished velocity differencebetween the flow along the pressure side and suction side of the blade,leading to a smaller secondary flow and thus lower losses, as well as byvirtue of a reduction in flow velocity in the rotor outlet, leading tolower losses in the following diffuser.

In this example, the blade outlets of the impellers additionally have aninclination γyb2 in the direction of rotation, as can be seen from FIG.4, with a view to reducing the tension increase in the blade outletwhich arises due to the increased blade angle βb2.

From FIG. 4 it can be seen that the diameter ratio D±/Du of the impellerbetween inlet and outlet lies within the range 0.50-0.62. In anotherillustrative embodiment, D±/Du can lie within the range 0.50-0.58.Turbocompressors available on the market have diameter ratios Di/Dubetween about 0.63 and 0.70. This diminution of the diameter ratioresults in an increase in the radius of curvature of the stream linesclosest to the blade tip. The flow along these stream lines, whichdiffuses, i.e. has a decreasing velocity (relative to the blades), isfavorably affected by the somewhat larger radius of curvature andthereby increased length and acquires, inter alia, a lesser tendency toseparate from the wall, so that the efficiency, above all at highpressure ratios, is improved.

In order to increase the pressure build-up, both compressors areprovided with guide vanes in the diffuser downstream of the respectiveimpeller. This diffuser is of the LSA (Low Solidity Airfoil) type, whichmeans that it is provided with aerodynamically configured blades, thelength of which has a ratio to the distance between the blades (pitch),along the periphery in the blade inlet, which lies within the range0.7-1.5. This diffuser type, unlike that which is used in gas turbinecompressors, i.e. with long diffuser blades, has the characteristic ofnot reducing the stable working range of the compressor at high pressureratios.

An outlet diffuser 37 is placed after the low-pressure turbine 21 inorder to extract dynamic pressure from the turbine. The diffuser opensout into an exhaust collector 38, which guides gases out to the exhaustpipe 24.

The high-pressure turbine 17 driving the high-pressure compressor 19 isof the radial type, having a turbine wheel which, for rotation atrelatively high rotation speeds, is constructed with small diameter.

As can be seen from FIGS. 5 and 6, which show characteristics for acompressor having a wheel diameter ratio D±/Dur blade outlet angle βb2and bladed diffuser in accordance with an aspect of the invention, incomparison to a compressor according to the state of the art asignificantly higher efficiency can be attained, above all at highpressure ratios, without any reduction in the stable operating range.Note that the rotation speed lines are not identical for the twocompressors, but that the diagrams only constitute an illustration ofthe efficiency and stable operating range for the two types ofcompressors.

An increased blade angle βb2 nevertheless results in a diminution of thepressure increase for a given rotation speed. In order to compensate forthis, a higher rotation speed or larger wheel diameter is required,which means that the tensions in the impeller, above all in the hub ofthe disk and the outlet of the blade, increase for the same pressureratio. An unexpected effect is however that, for blade angles βb2 aboveabout 40°, the optimal rotation speed for the compressor constructionincreases more than is required to maintain the pressure increase, whichmeans that the diameter can therefore even be diminished. This can beseen from FIGS. 7 and 8, in which FIG. 7 shows the work factor δh0/U² asa function of the blade angle βb2, where δh0 is as a single figure piincrease and U is the peripheral velocity of the impeller. An increasein the blade angle βb2 from, for example, 40° to 50° means that the workfactor diminishes by about 6%. In order to maintain the pressure ratio,the peripheral velocity U must then be increased by about 3%(√1.06=1.03), assuming unchanged efficiency. Optimal rotation speed canbe read off from Diagram 4, which shows efficiency as a function ofspecific rotation speed Ns and blade angle βb2. Specific rotation speedNs is here defined as Ns=ω*√V/H_(ad))^(3/4), where ω=angular velocity,V=inlet volume flow, Had=adiabatic single-figure pi increase(=Cp*T_(0,in)*((pressure ratio)^(((K−1)K))−1)). From Diagram 4 it can beseen that optimal Ns and thus rotation speed, given unchanged volumeflow, pressure ratio and inlet state, increases by about 4% when theblade angle βb2 is increased from 40° to 50°. The turbine which willdrive the compressor can be reduced in diameter at least equivalently tothe higher rotation speed of the compressor, which means low polarmoment of inertia.

The invention should not be deemed to be limited to the above-describedillustrative embodiments, but rather a host of further variants andmodifications are conceivable within the scope of the following patentclaims. For example, the turbocharger unit according to an aspect of theinvention is described in connection with a six-cylinder diesel enginewith two-stage turbocharge, but the invention is applicable to a fullvariety of piston engines from one cylinder and upward, and which aredriven in two or four-stroke with one-stage as well as two-stageturbocharge. Aspects of the invention can also be applied to ship'sengines and to engines having other cylinder capacities than theaforementioned. In the case of a two-stage turbocharge, thehigh-pressure turbine 17 can be without inlet guide vanes, oralternatively can be provided with fixed or geometrically rotatableinlet guide vanes 17 a, and the low-pressure turbine 21 can be of theradial type as well as of the axial type.

1. A turbocharger unit for an internal combustion engine having at leastone exhaust line for the evacuation of exhaust gases from the combustionchamber of the engine and at least one inlet line for the supply of airto the combustion chamber, comprising a turbine, which interacts with acompressor in order to extract energy from the exhaust-gas flow of theengine and pressurize the inlet air of the engine, wherein thecompressor is of the radial type and is provided with an impeller havingbackswept blades in which the blade angle, between an imaginaryextension of the center line of the blade between root section and tipsection in the direction of the outlet tangent and a line connecting thecenter axis of the impeller to the outer tip of the blade, is at leastabout 40 degrees, and the ratio between the inlet diameter of theimpeller and its outlet diameter lies within the range 0.50-0.62, and isfurther provided with a diffuser, which is provided with blades, thelength of which has a ratio to the distance between them, along theperiphery in the blade inlet, within the range 0.7-1.5, for diminishingvelocity difference between the flow along a pressure side and a suctionside of the blade for increasing compressor efficiency.
 2. Theturbocharger unit as claimed in claim 1, wherein the blade angle is atleast about 45°,
 3. The turbocharger unit as claimed in claim 1 whereinthe blade inclines, at an angle of inclination greater than 0 degrees,in the direction of rotation in the outlet of the impeller.
 4. Theturbocharger unit as claimed in claim 1 wherein the angle of inclinationis at least about 30 degrees.
 5. The turbocharger unit as claimed inclaim 1 wherein the impeller is machine-worked from an aluminum alloyforging.
 6. The turbocharger unit as claimed in claim 1 wherein theimpeller is machine-worked from a titanium alloy.